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1 THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS -345 E. 47th St., New York, N.Y. 117 The Society shall not be responsible for statements or opinions advanced in papers or discussion at meetings of the Society or of its Divisions or Sections, or printed in its publications. Discussion is printed only If the paper is published in an ASME Journal. Papers are available from ASME for 15 months after the meeting. Printed in U.S.A. Copyright 1994 by ASME 94-G1-73 DEVELOPMENT OF CENTRIFUGAL COMPRESSOR FOR 1 kw AUTOMOTIVE CERAMIC GAS TURBINE Hiroshi Uchida, 1 Mutsuo Shiraki, 1 Akinobu Bessho,1 and Yolchi Yagi2 Japan Automobile Research Institute, Incorporated Tsukuba, lbaraki, Japan V ABSTRACT In Japan, a program of research and development of a 1 kw automotive ceramic gas turbine (CGT) has been carried out in the Petroleum Energy Center with active cooperation of petroleum, automobile and ceramics industries as well as other related industries. As a part of this research and development program, we have studied and developed a centrifugal compressor with variable inlet guide vanes for COT engines. There has been a strong demand for a compressor with a high efficiency and a wide flow range. The compressor performance goals are an adiabatic efficiency of 81% and a surge margin of 8% under maximum power operating conditions. This paper describes the methods for designing impellers, diffusers and variable inlet guide vanes, and presents the results of compressor performance tests. The test results reveal that the surge margin and compressor efficiency at partial load are improved by using inlet guide vanes. NOMENCLATURE absolute velocity peripheral velocity relative velocity diameter tas, enthalpy increment through compressor stage a av air flow angle (relative to meridional) vane setting angle of VIGVs work factor 'Toyota Central R&D Labs., Inc. 'Nissan Motor Co., Ltd. SUBSCRIPTS compressor inlet 1 impeller inlet 2 impeller exit 3 compressor owlet mean or meridional peripheral relative INTRODUCTION A seven-year program for developing a 1 kw automotive ceramic gas turbine (COT) was started in 199. The aims of the program are to design and manufacture a 1 kw automotive COT with excellent exhaust gas emissions and thermal efficiency of not less than 4%, and to prove its superior potential by means of bench tests. The COT engine targeted is of a single shaft type as shown in Figure 1. It is composed of a centrifugal compressor with variable inlet guide vanes (VIGVs), a radial turbine, a can type combustion chamber and two regenerative rotary heat exchangers. The pressure ratio under maximum power operating conditions is five, and the maximum temperature at a turbine inlet is 135t. The aerodynamic design and basic structure design of a centrifugal compressor with VIGVs were carried out. The aerodynamic characteristics were evaluated by the following 1/4 tests. (1) The test of the VIGVs system (2) The test in the performance of a compressor without VIGVs (3) The test in the performance of a compressor with VIGVs The characteristics of the flow at the location corresponding to an impeller inlet were measured by the test of the VIGVs system. The compressor efficiency and its flow range were Presented at the International Gas Turbine and Aeroengine Congress and Exposition The Hague, Netherlands - June 13-16, 1994 Downloaded From: on 7/2/218 Terms of Use:

2 COMBUSTION CHAMBER COMPRESSOR IMPELLER VARIABLE INLET GUIDE VANES TURBINE ROTOR REGENERATOR FIG. 1 CGT ENGINE STRUCTURE evaluated by the performance of compressors both without the VIGVs and with the VIGVs. This paper describes mainly the results of the compressor performance tests. PERFORMANCE GOALS AND DESIGN SPECIFICATIONS The performance goals and basic design specifications of the compressor are shown in Table 1. The prewhirl angle requires the VIGVs to range from -1 to 6 degrees. The rotational speed at a design point is 11, rpm. This rotational speed is determined so as to secure both high aerodynamic performance and strength reliability of a ceramic turbine rotor. Accordingly, the impeller of the compressor has to be of small size. Also, the relative Mach number at the inducer tip exceeds unity because the specific speed is set at a higher value. Under these conditions, an adiabatic total-to-total efficiency of 81% and a surge margin of 8% have to be achieved. In order to achieve these goals, the following points need to be investigated. (1) To make the blade thickness as small as possible (2) To make the back swept blade angle as large as possible (3) To make the tip clearance as small as possible (4) To reduce the Mach number at the inducer tip by means of prewhirl with VIGVs For automotive use, it is also important to make the flow range shifted by VIGVs as wide as possible. TABLE 1 COMPRESSOR SPECIFICATIONS AND PERFORMANCE GOALS Type Prewhirl Angle Flow Rate Turbine Rotor Single Stage Centrifugal (with VIGVe) -1-6 degrees.445 kg/sec Stage Pressure Ratio 5 Rotational Speed Target Adiabatic Efficiency (total-to-total) 11, rpm 81 % (79 %: intermediate) Target Surge Margin 8% COMPRESSOR VIGV SYSTEM Impeller VIGV Scroll Diffuser Shroud FIG. 2 STRUCTURE OF THE COMPRESSOR SYSTEM AERODYNAMIC AND MECHANICAL DESIGN The structure of the compressor developed is shown in Figure 2. This compressor is composed of VIGVs, a backward swept centrifugal impeller, a vane diffuser and a scroll. The scroll has two exits leading to two heat exchangers. VIGV Design The structure of the radial type VIGVs is shown in Figure 3. The VIGV has an articulated design (with fixed leading edges, and movable trailing edges) using a modified NASA series thickness distribution. The number of vanes is 19. The vane cord length and the vane width were determined through one-dimensional flow analysis to be 25 mm and 21 mm, respectively, so that the prewhirl angle from -1 to 6 degrees can be obtained at the impeller inlet. The two intake ducts are installed upstream so that a uniform circumferential distribution of flow can be obtained. The VIGVs are moved by a hydraulic actuator with links. 2 Downloaded From: on 7/2/218 Terms of Use:

3 D-II 31.5 Prewhirl Angle : -1 : 6 FIG. 3 VIGV STRUCTURE Impeller Design A backward swept centrifugal impeller having 1 full blades and 1 splitter blades was designed for the COT compressor. Its exit blade angle in the radial direction was enlarged as much as possible in order to obtain the wide flow range. The blade thickness of the impeller was studied so that it was made as small as possible through the elastic threedimensional stress and frequency analysis. The blade curvature, the number of blades and the blade thickness distribution were examined so that a smooth velocity change and uniform blade load were obtained through three-dimensional flow analysis. In 199, the first design (hereafter called 1)-I) impeller was manufactured and its performance was evaluated by the compressor performance test. The test results showed that the adiabatic efficiency of a compressor with the D-I impeller was lower than the intermediate target efficiency of 79% and that the flow range of the compressor was very narrow. In 1992, the second design (hereafter called D-H) impeller was designed and manufactured. The results of the compressor performance tests showed that both the efficiency and the flow range were improved. The D-11 impeller design features are compared to those of the impeller below. Figure 4 shows the geometries both of the D-I and D-H impellers. The D-II impeller was obtained by modifying the D-I impeller. The impeller exit diameter was modified from 95 mm to 14 mm and the average exit blade angle in the radial direction was modified from 35 degrees to 4 degrees. In addition, the exit blade width was modified from 4.6 mm to 4.3 mm so that the exit flow area might be the same. Figure 5 shows the blade angle distribution on the shroud and the hub surfaces along the meridional length. The inlet blade angle on the hub surface of the D-11 impeller is larger than that of the D-I impeller. Figure 6 shows the normal blade thickness distribution on the shroud and the hub surfaces. The 13-I and D-11 impellers have the same blade thickness distribution. On the shroud surface, the blade thickness is an almost constant value of.4 mm. The blade thickness on the hub surface has a maximum value of 2.1 mm at 4% of the meridional length,.6 mm at the leading edge and 1.2 mm at the exit. 1 FIG. 4 IMPELLER CONFIGURATION HUB.. -,./ SHROUD DI : : D 2 HUB C,, ` ' SHROUD Meridional Distance (nondimensional) FIG. 5 IMPELLER BLADE ANGLE DISTRIBUTION 3 Downloaded From: on 7/2/218 Terms of Use:

4 2.1 Maximum Slress 7 &Pa Axis Side 27E.1 4E.1 367E Meridional Distance (nondimensional) FIG. 6 IMPELLER BLADE THICKNESS DISTRIBUTION Second Design 2 8E:31 1 3E Figure 7 shows the stress distribution of the D-11 impeller calculated under the conditions of the material of Ti-6A1-4V and a rotational speed of 11, rpm. The impeller stress has a maximum value of 714Pa near the center of the axis. The maximum stress is smaller than the material strength of 98 MPa (473K). Figure 8 shows the Campbell diagram obtained by the frequency analysis. The first vibratory mode interferes with the VIGVs excitation at a rotating speed of 29, rpm. This rotating speed is lower than the idle speed of a COT engine. It is thought that the possibility of damage due to the VIGV excitation is limited, considering the long flow path from the VIGV to the impeller inlet. piffuser Design The single-stage radial diffuser with 25 vanes was designed considering the size of the compressor so as to make it as small as possible. The diffuser vane has a straight camber line and has been designed using the NASA65-A6 series thickness distribution modified to a maximum thickness of 1.1 tom. The diffuser inlet diameter has the value 1.1 times as large as the impeller exit diameter for the D -I and diffusers. The diffuser vane width has the value.9 times the impeller exit blade width. The inlet vane angle and throat width of the D-II diffuser were determined so that its throat area might be the same as that of the D-I diffuser. The material of the diffusers is SUS34. The geometric design summary both of the impeller and the diffuser is shown in Table 2. Also, pictures of the D-I and D-11 impellers with diffusers are shown in Figure 9. AERODYNAMIC PERFORMANCE TESTS Performance Test of the VIGVs The VIGVs test rig is shown in Figure 1. The air is FIG. 7 IMPELLER STRESS DISTRIBUTION 3 _ D-II Impeller 42. / te -.. 4p," ty yet; es.6e," c.,,. c -, )... i.,. &.- f...- \c> 3rd Mode rte/ NC) nd Mode /," 1st Mode,. a or I I X1 4 Rotational Speed (rpm) FIG. 8 TITANIUM IMPELLER CAMPBELL DIAGRAM breathed by means of a centrifugal blower which was set up downstream. Its flow rate is measured using the orifice located downstream. The distributions of the velocity, the total pressure and the prewhirl angle were measured using a Pitot probe at the location corresponding to the impeller inlet. This Pitot probe with three holes is able to be traversed between the shroud and the hub surfaces. The performance tests for VIGVs were carried out and objectionable acoustic noise was identified when the VIGVs were closed to 4 degrees or higher and when the air flow rate was higher than.18 kg/sec as shown in Figure 11. This phenomenon was 4 Downloaded From: on 7/2/218 Terms of Use:

5 TABLE 2 IMPELLER AND DIFFUSER GEOMETRIES tr LAJ -. First Design Second Design WM Tip Diameter (mm) Het Hub Diameter (mm) edt Diameter (mm) Exit Blade Width (mm) Traversed Location. Impeller Leading Edge sir, VIGVs X Inlet Tip Blade Angle (degree) Average Exit Blade Angle (degree) Blade Number 1 / 1 1 / 1 Vane Inlet Diameter (mm) Vane Outlet Diameter (nn) DIFFUSER Vane Width (mm) Inlet Vane Angle (degree) Vane Number Throat Width (mm) Throat Area (cm2) FIG. 1 VIGV TEST RIG ;calialtitatwwssm:wirp,a; Corrected Flow Rate (kg/sec).5.design Flow Rate.4 Domain with Noise.3.2 (;) without Noise.1 with Noise Vane Setting Angle: a (degree) FIG. 9 IMPELLERS AND DIFFUSERS FIG. 11 DOMAIN OF THE VIGV ACOUSTIC NOISE similar to the experience on AGT11, and this problem was resolved by installing two flow tabs in the VIGVs as shown in Figure 12. The performance tests for the VIGVs with the flow tabs were carried out. Figure 13 shows the prewhirl angle measured at the location according to the impeller inlet mean radius. The reverse prewhirl angle of -5 degrees was observed for a vane setting angle of degree. The prewhirl angle of 59 degrees was obtained for a vane setting angle of 6 degrees. rmance Test for the Comoressor without The compressor performance test rig without VIGVs is shown in Figure 14. The test rig has a radial turbine to drive the compressor and has two shafts connected with each other by a spline coupling. The compressor intake duct is straight and has a bellmouth to measure the air flow rate. The impeller rotating speed and the oscillating displacements of the two shafts were measured using three gap-sensors located near the shafts. The compressor performance tests were carried out under the condition that the driving turbine inlet temperature was kept constant at 873 K. The compressor adiabatic efficiency (total-to-total) was evaluated using the ambient temperature and the ambient pressure measured at the bellmouth inlet and the total temperature and the static pressure measured on the scroll outlet ducts. The compressor exit static pressure was converted to the total pressure through 1-D continuity. Furthermore, the wall static pressures both on the shroud and hub surfaces were measured at the impeller inlet and the diffuser inlet, and these pressures were used for evaluation of the intake duct pressure loss, the impeller 5 Downloaded From: on 7/2/218 Terms of Use:

6 Flow Tahq COMPRESSOR TURBINE ftir A Diffuser e er Gap Sensor S line Co Gap Sen sor 44 1/ 1 i. NIElitscl, ' PO.'ilii, ,41 it, --- ir.,..z,:a.:..:---...,, _ ::...- a Thilem=4-11.4s It ga-relatatari tr ii I le- - "Illi - drawn Thlwrip,er 6 1 P1. 12 VIGV FLOW TABS Vane Setting Angle: av (degree) FIG. 13 PREWHIRL ANGLE (VIGV TEST) performance and the diffuser performance in the onedimensional flow analysis. On the other hand, the surge flow rate was experimentally obtained by identifying the surge phenomena such as objectionable periodic noise and vibration. Figure 15 shows the characteristics of the D-I and D-I1 compressors in comparison with each other. The performance of the D-TI compressor was significantly improved compared to that of the D-I compressor. The adiabatic efficiency of the D-II compressor increased by 2.5% and reached the value of 79.5%, which was hi gher than the FIG. 14 COMPRESSOR PERFORMANCE TEST RIG WITHOUT VIGVS.8 c , satiaintisiamak*ii num I:. aiiuuuiiinui 4.) na. MM. MSS: Improved Region, t D-11 rc:.. D-1 aid # V '' : I 818 o l I J Corrected Flow Rate (kg/sec) FIG. 15 CHARACTERISTICS OF COMPRESSOR WITHOUT VIGVS Point intermediate target value of 79% at the design point (a flow rate of.445 kg/sec and a pressure ratio of 5). Moreover, the wide flow range was obtained due to the significant improvement of the surge characteristics. As the results, the surge margin increased to 7% and reached near 6 Downloaded From: on 7/2/218 Terms of Use:

7 1.7% Second Design (-1:1), at 1, rpm First Design (D-ll, at 11, rpm Impeller Exit.5 1 Design Flow Rate.5 2 [ Corrected Flow Rate) 2( kg/sec ) 2 FIG. 16 TOTAL PRESSURE LOSS IN INTAKE DUCT (WITHOUT VIGVS) 1. 2 ( 54 nvsec ) Inducer Leading Edge (Mean) u, C 1 ( 17 rrvsec ) C: Absolute Velocity W: Relative Velocity U: Peripheral Velocity the desired target value of 8%. In the case of the D-II compressor, however, the design pressure ratio of 5 was achieved at 1, rpm, which was lower than the design speed of 11, rpm due to the increment of the impeller diameter. Figure 16 shows the total pressure loss in the intake duct which was obtained from the results of the static pressure measurements at the impeller inlet. The pressure loss of 1.7% is indicated at the design flow rate. The velocity diagrams at the inducer leading edge and the impeller exit are shown in Figure 17. The impeller exit peripheral velocity is 54 m/sec at a rotational speed of 1, rpm in the case of the 13-11I compressor and 11, rpm in the case of the D-I compressor. The impeller inlet relative velocity and impeller exit absolute velocity in the D- I! compressor are lower than those in the D-I compressor. Accordingly, it is thought that the adiabatic efficiency and the surge margin of the D-rs compressor are improved due to the decrements both of the impeller inlet relative Mach number and the diffuser inlet Mach number. Test In Performance of Compressor with VIGVs The test rig with VIGVs is shown in Figure 18. The wall static pressures at the VIGV inlet and the impeller inlet and outlet were measured in order to evaluate the total pressure losses both in the intake duct and in the VIGVs and the performances of an impeller and a diffuser. The performance tests of the compressor with VIGVs having two flow tabs were carried out with four VIGV vane setting angles:, 2, 4 and 6 degrees. Furthermore, the effects of the VIGV vane setting angle on the compressor performance were evaluated and analyzed through one-dimensional flow analysis. Figure 19 shows the effects of the VIGV vane setting angle on the characteristics of the D-II compressor. The inlet total pressure was obtained by converting the VIGVs inlet static pressure through 1-D continuity. The total pressure loss between the bellmouth inlet and the FIG. 17 IMPELLER VELOCITY DIAGRAMS AT DESIGN POINT VIGVs inlet is 2 % at the design flow rate. The pressure ratio when av= degree is higher than that of the compressor without VIGVs due to the reverse prewhirl. flow shown in Figure 13. The adiabatic efficiency was significantly improved from 79% to 81.5% as av was shifted from degree to 2 degrees at 1, rpm. However, the surge flow rates are similar to each other. Besides, as ccv was shifted to 4 and 6 degrees, the surge flow rate decreased step by step at the rotational speeds higher than 7, rpm. On the contrary, the surge flow rate at the rotational speeds lower than 6, rpm was not improved. The impeller inlet prewhirl angles at were evaluated for the four cases through the one-dimensional flow analysis using the angular momentum theory. The nomenclature in the velocity diagrams is shown in Figure 2. The compressor work factor X is expressed by the following equation. - HiCut EdHco: enthalpy increment] Accordingly, the impeller inlet circumferential velocity Cm is obtained using the following equation. cut 152 Dicr/Dz [Did mean diameter at the impeller inlet] In equation (2), X* indicates the work factor of the compressor without VIGVs (Cur4). Figure 21 shows the work factors for ccvd,/, 2,4 and 6 (1) (2) 7 Downloaded From: on 7/2/218 Terms of Use:

8 Atomospheric Pressure (1 point) Atomospheric Temperature (6) VIGVs Inlet Pressure (12) Compressor Outlet Temperature (12) Pressure (6) Impeller Outlet Pressure Shroud:12, Hub23 Impeller Inlet Pressure (Shroud:12, Hub:1 NTAKE DUCT BELLMOUTH FIG. 18 COMPRESSOR PERFORMANCE TEST RIG WITH VIGVS D-ft Compressor.8 >, 2.7 : iiiiiilitil M11111=11 iiimhjiiiiti i Pressure Ratio 5 4 Vane Setting c oangle - 2 s Ai II 3 to ra! % 111 L I i I i I I q A ' i A 1 t Corrected Flow Rate (kg/sec) FIG. 19 CHARACTERISTICS OF COMPRESSOR WITH VIGVS FIG. 2 NOMENCLATURE OF VELOCITY DIAGRAMS degrees compared to that of the compressor without VIGVs. The work factor when av= degree is higher than the value of X*, and it indicates the reverse prewhirl for ay of degree. The prewhirl angles obtained by using equation (2) are shown in Figure 22. These prewhirl angles are lower than those obtained by the VIGV test shown in Figure 13. The impeller inlet velocity distributions are obtained using the analytical prewhirl angles and the impeller inlet static pressure measurements both on the shroud and hub. Figure 23 shows the impeller inlet velocity diagrams at a surge point of 1, rpm. The absolute velocity at the 8 Downloaded From: on 7/2/218 Terms of Use:

9 .8 X* %.. % A.* : without VIGVs 6 = 6 Itif Xqz 4 GA *ea. av = 4 taserahathatatto Flow Coeficient : tp2 (=Cm 2 /U -2 MI) 4.8%) = 2 ocen1/44) ce, = Neat enolirb Ps 114%41., Corrected Flow Rate (kg/sec) FIG. 21 WORK FACTORS inducer tip decreases as the vane setting angle increases. On the contrary, the absolute velocity at the inducer hub increases as the vane setting angle increases. As a result, the increment of the circumferential velocity at the inducer tip due to the increment of the vane setting angle is less than that at the inducer hub. Accordingly, it is thought that the circumferential velocity at the inducer tip must increase in order to decrease the surge flow rate still more. CONCLUSIONS (1) The adiabatic efficiency and the surge characteristics were significantly improved by increasing the impeller tip diameter and the impeller exit blade angle. As a result, an adiabatic efficiency of 79% and a surge margin of 7% near the intermediate target value were achieved. However, design pressure ratio of 5 is obtained at 1, rpm which is lower than design speed of 11, rpm. (2) The adiabatic efficiency was increased from 79% to 81.5% at 1, rpm by shifting the VIGV vane setting angle from degree to 2 degrees. However, the surge line was not improved. (3) The surge flow rate decreased step by step as the VIGV vane angle was shifted to 4 and 6 degrees at rotational speeds higher than 7, rpm. The surge line did not shift at the rotational speeds lower than 6, rpm. ACKNOWLEDGMENTS The authors are grateful to the Agency of Natural Resources and Energy, the Ministry of International Trade and Industry, for making this development possible and to the program management at the Petroleum Energy Center. Also, we would like to thank members of the CGT organization in Japan Automobile Research Institute for their technical support FIG. 22 PREWHIRL ANGLES (COMPRESSOR PERFORMANCE TEST) at Surge Point (1, rpm).aw - : =2. -4 : =6 FIG. 23 INDUCER LEADING EDGE VELOCITY DIAGRAMS REFERENCES T. Itoh and H. Kimura, 1992, "Status of the Automobile Ceramic Gas Turbine Development Program," ASME Paper 92-GT-2. NASA CR-17518, 1984, Advanced Gas Turbine Technology Project 1984 Annual Report. NASA CR , 1985, Advanced Gas Turbine Technology Project 1985 Annual Report. NASA CR , 1985,Advanced Gas Turbine (AGT) Technology Development Project 1985 Annual Report. NASA CR-18891, 1987, Advanced Gas Turbine (AC!') Technology Development Project Final Report, Garrett Auxiliary Power Div. 9 Downloaded From: on 7/2/218 Terms of Use: